Hydrodynamic bearing

ABSTRACT

A rotatable shaft/thrust plate combination is disposed within a sleeve to form a first clearance space between the shaft and the sleeve and a second clearance space between the thrust plate and the sleeve. The external faces of the thrust plate are exposed to air. The clearance spaces are filled with a liquid lubricant and the sleeve includes pressure equalization ports connecting the first and second clearance spaces. Surface tension dynamic seals are provided between axially extending surfaces of the thrust plate and sleeve. The equalization ports balance the hydrodynamic pressures in the lubricant to prevent the lubricant being pumped through one of the dynamic seals. The resulting bearing provides high precision with low repetitive and nonrepetitive runouts. The bearing provides hydrodynamic support of both radial and axial loads and the bearing seal is relatively insensitive to orientation of the spindle and minimizes the generation of debris and contaminating particles.

BACKGROUND OF THE INVENTION

This invention relates to precision hydrodynamic bearings.

One important limitation to increasing track density of computer diskdrives is spindle bearing performance. A disk drive whose spindlebearing has low runout can accommodate higher track densities whichresults in more data storage capacity per disk.

The kinematics of the spin axis of a spindle bearing determine theprecision of the bearing. As the journal spins relative to the sleeve,the spin axis may trace out a path or orbit. The motion of this axistypically has components that are synchronous with the spin andrepetitive in nature. These motions are termed repetitive runout. Othercomponents of spin axis motion may be asynchronous and nonrepetitivewith respect to spin. These components are termed nonrepetitive runout.As a general rule, spindle bearing precision is increased as repetitiveand nonrepetitive runouts are decreased.

Ball bearing spindle systems make up the majority of prior art diskdrives. The kinematics of the rolling elements in ball bearings resultin relatively large nonrepetitive runout. This results from the factthat the lubricant film thicknesses in ball bearings are very thinproviding little attenuation of geometric defects in the bearing. Inaddition, ball bearings produce forces on the disk drive structure towhich it is attached which are of relatively high frequency and largeamplitude.

Hydrodynamic spindle bearing designs are also known. The Hewlett PackardModel No. 9154A, 3.5 inch micro-Winchester disk drive incorporates ahybrid hydrodynamic ball bearing spindle. The performance of thisbearing is degraded by the incorporation of the ball bearings. ThePhillips video 2000 videocassette recorder utilizes a hydrodynamicbearing which employs grease as the lubricant limiting operation to lowspeeds. Other known hydrodynamic spindle bearings for disk drives employa ferromagnetic fluid as the lubricant for the bearing. This fluid isretained or sealed in the bearing by magnetic fields set up in polepieces at each end of the bearing. Unless the magnetic fields andclearances are very precisely matched at each end of the bearing, oneseal will be stronger than the other and when the bearing heats up, thelubricant can be spilled. See U.S. Pat. No. 4,526,484.

SUMMARY OF THE INVENTION

In general, the hydrodynamic bearing according to the instant inventionincludes a rotatable shaft/thrust plate combination disposed within asleeve forming a first clearance space between the shaft and the sleeveand a second clearance space between the thrust plate and the sleeve.The external faces of the thrust plate are exposed to air and theclearance spaces are filled with a liquid lubricant. The sleeve includespressure equalization ports connecting the first clearance space and thesecond clearance space. In a preferred embodiment, the bearing includessurface tension dynamic seals between axially extending surfaces of thethrust plate and sleeve. These axially extending surfaces of the thrustplate and sleeve diverge toward the ends of the bearing to form thedynamic seal. The divergence may be a straight taper having an angle ofapproximately 2°. The pressure equalization ports include axiallyextending passageways in communication with radially extendingpassageways to connect the first and second clearance spaces. Theradially extending passageways may be located near the center of thebearing. The bearing may also include relief patterns in opposedsleeve/thrust plate faces to generate inwardly directed radial forces.

In one embodiment of the invention, the bearing includes a cylindricalsleeve including a portion having a smaller inside diameter. A shaftincluding a portion having a diameter adapted to form a first clearancespace with respect to the smaller diameter portion of the sleeve fitswithin the sleeve. A pair of thrust plates are disposed on the shaft toform second clearance spaces with respect to radially extending faces ofthe smaller diameter portion of the sleeve, the external faces of thethrust plate being exposed to the air. The clearance spaces are filledwith a liquid lubricant. The smaller diameter portion of the sleeveincludes plural axially extending passageways in liquid communicationwith radially extending passageways interconnecting the first and secondclearance spaces. Surface tension seals are provided between the thrustplates and sleeve.

Another aspect of the invention is a method for introducing lubricantinto the hydrodynamic bearing to avoid incorporating air. The bearing isplaced in a vacuum chamber above a liquid lubricant and the chamber isevacuated to a pressure below atmospheric pressure. The bearing issubmerged into the lubricant and the pressure in the chamber is raisedto atmospheric pressure which forces the lubricant into the clearancespaces in the bearing. After the bearing is filled, it can be exposed toultrasonic energy to expel any residual air. The vacuum chamber can alsobe repeatedly cycled between a high and a low pressure to expel residualair.

In another, particularly preferred embodiment of the invention, thebearing incorporates both external and internal surface tension seals ateach end of the bearing. In this embodiment, there is an air spacebetween the two ends of the bearing. This embodiment results in areduced evaporation rate from the seals, improved moment stiffness, andfaster thermal transient response.

In yet another aspect of the invention, the shaft and sleeve includemating tapered portions at each end of the bearing defining lubricantfilled clearance spaces for supporting radial and axial loads. Eachclearance space is sealed by an internal and an external surface tensiondynamic seal and pressure equalization ports are provided to connect theinternal and external seals. In this embodiment, the shaft is acontinuous unit without a separate thrust plate portion. No O-ring sealsare required.

The hydrodynamic bearing of the instant invention achieves lower levelsof runout than ball bearings as a result of a thick film of lubricantwhich separates the sliding metal surfaces. This film provides a highdegree of viscous damping which significantly attenuates nonrepetitiverunout to levels which are less than state of the art rolling elementbearings. In addition, the bearing generates forces on the structureattached to it which are low frequency and low amplitude relative toball bearings. This reduction in the forcing function bandwidth andamplitude minimizes other vibrations in the disk drive and furtherimproves tracking performance. The pressure equalization ports reducepressure differentials which are caused by pumping actions inside thebearing. Because of the pressure balancing, the bearing does not tend topump lubricant in a preferential manner through one seal or the other.Thus, only the external pressure differential across the bearinginfluences the position of the dynamic seal interfaces. The surfacetension seals of the present invention do not leak nor do they generatesolid debris.

BRIEF DESCRIPTION OF THE DRAWING

FIG. 1 is a cross sectional view of the bearing of the invention;

FIG. 2 is an elevational view of the sleeve portion of the bearing;

FIG. 3 is an expanded view of a portion of FIG. 1;

FIG. 4 is an expanded view of a portion of FIG. 1;

FIG. 5 is an expanded view of a portion of FIG. 4;

FIG. 6 is a schematic illustration of the method of filling the bearingwith lubricant;

FIG. 7 is a cross sectional view of a particularly preferred embodimentof the present invention; and

FIG. 8 is a cross sectional view of an embodiment of the inventionutilizing a tapered shaft.

DESCRIPTION OF THE PREFERRED EMBODIMENT

A hydrodynamic bearing 10 shown in FIG. 1 includes a sleeve 12 includinga portion of smaller inside diameter 14. A journal or shaft 16 fitswithin the sleeve 12 forming a first clearance space 18. The journal 16may include a recess 20. Thrust plates 22 and 24 rest on the journal 16and are sealed by means of O-ring seals 26. The thrust plates 22 and 24form second clearance spaces 28 with respect to radially extendingsurfaces of the smaller inside diameter portion 14 of the sleeve 12. Theportion 14 of the sleeve 12 also includes axially extending passageways30 and radially extending passageways 32. As shown in FIG. 2 thepassageways 30 and 32 are arranged around the circumference of thesleeve 12. Four sets of passageways 30 and 32 are shown in FIG. 2 butmore or fewer may be employed. FIG. 2 also shows spiral relief patterns34. These relief patterns cooperate with patterns on the journal togenerate radially directed inward hydrodynamic pressure.

Relative rotation between the journal 16 and the sleeve 12 is providedfor by the clearance spaces 18 and 28. Suitable dimensions for theclearance spaces 18 and 28 are 0.0002 to 0.001 inches and 0.0005 to0.002 inches, respectively. These clearance spaces are filled with alubricant such as oil which reduces wear between the journal and sleeveand provides a medium through which a hydrodynamic pressure field may begenerated. Relative rotation or radial motion between the journal 16 andsleeve 12 is required to set up the hydrodynamic pressure field. Thehydrodynamic bearing 10 supports loads by metal to metal contact whenthere is no relative motion. During normal operation, the spinning ofthe journal 16 sets up a steady pressure field around the clearancespaces which pushes the journal and sleeve apart and thus prevents metalto-metal contact. The hydrodynamically pressurized film provides thestiffness needed to support the radial load of the disk, motor andassociated hardware. Note that the hydrodynamic film stiffness is ameasure of the resistance of the clearance space to change size underthe influence of a load.

Axial loads along the journal 16 spin axis are supported by thehydrodynamic pressure field in the clearance spaces 28 between thethrust plate faces and the sleeve portion 14. The amount of separationbetween the thrust plate faces and sleeve is controlled by thehydrodynamic film stiffness and the applied axial load (usually theweight of the entire rotating assembly). Pressure building geometriessuch as the relief pattern 34 shown in FIG. 2 are employed to generatefilm stiffness of sufficient magnitude.

The sealing of the lubricant within the hydrdynamic bearing 10 will nowbe described in conjunction with FIGS. 1, 4 and 5. There are two typesof seals in the bearing 10, namely, static and dynamic seals. Staticseals 26 which are preferably O-ring seals prevent lubricant leakagebetween the thrust plates 22 and 24 and the journal 16. They are calledstatic seals in that there is no relative rotation or sliding betweenthe thrust plates 22 and 24 and the journal 16. Dynamic sealing isrequired in the clearance space 36 between the thrust plates and thesleeve. These seals must not leak or generate solid debris. Sealing isprovided by surface tension-capillary seals in which a lubricant-airinterface 38 provides the surfaces forces.

As shown in FIG. 5, two components, the liquid-gas (lubricant-air)interface 38 and the solid surfaces of the thrust plates and sleeve makeup each seal. Surface tension forces directed axially away from each endof the bearing indicated by the arrows 40 balance the forces due topressure differentials which may be applied across each interface asindicated by the arrows 42 and a force due to gravity. The magnitude ofthe axial surface tension forces depends on the wetted perimeter of theliquid-gas interface 38, the surface tension (a property of the liquidlubricant), the taper angle and the contact angle. The forces due topressure differentials are dependent on the pressure differentials andthe lubricant-air interface area. Since the solid boundaries of the sealare tapered, the wetted perimeter and area of the interface vary withthe axial position of the interface. As a result, the axial position ofthe interface varies with pressure differences applied to the bearinguntil the surface tension forces and pressure forces balance. Stabilityof the interface is sensitive to the angle of taper. A taper angle ofapproximately 2° has been experimentally determined to be optimum forinsuring interface stability.

During bearing 10 operation, it is necessary that the pressures benearly the same at the lubricant side of each lubricant air interface38. This pressure balance is provided by the pressure equalization ports30 and 32 which connect the clearance spaces 18 and 28. Without theequalization ports, pumping actions inside the bearing may set uppressure differentials. For example, the thrust plates 22 and 24 producean inwardly directed radial pumping action. The equalization ports tendto equalize the pressures. Furthermore, the passages should maintain aconstant radial position in the neighborhood of the thrust plates. Thisrequirement prevents large pressure gradients from developing in thepassages due to the centrifugal pumping effects caused by the thrustplates. The bearing 10 is thus pressure balanced and does not tend topump the lubricant in a preferential manner through one seal or theother. Only the external pressure differential across the bearing,therefore, influences the position of the interfaces. The equalizationports coupled with the surface tension dynamic seals result in ahydrodynamic bearing of higher precision with respect to runout relativeto conventional bearing designs.

Lubricant must be introduced into the bearing in such way that a minimalamount of air is trapped in the bearing. This is necessary becausetrapped air in the bearing expands as the bearing heats up and tends topush the lubricant out of the bearing. A method for filling the bearingwith lubricant so as to minimize the amount of trapped air will bedescribed in conjunction with FIG. 6. First of all, the bearing 10 isplaced within a vacuum chamber 50 above the level of a liquid lubricant52. The vacuum chamber 50 is then evacuated to a suitable pressure belowatmospheric such as 5μ of mercury. The bearing 10 is then submergedwithin the lubricant 52, after which the pressure in the chamber 50 isallowed to rise to atmospheric pressure. As the pressure rises,lubricant is forced into the bearing through the clearance spacesbetween the thrust plates and sleeve. Residual air bubbles in thebearing may be removed by applying ultrasonic energy to the chamber 50within an ultrasonic tank 54. If necessary, additional residual air maybe removed by repeatedly cycling of the pressure in the chamber 50between a high and a low pressure.

FIG. 7 is a particularly preferred embodiment of the invention havingseveral advantages as compared to the embodiment of FIG. 1. A bearing 70includes a shaft 72 with thrust plates 74 and 76. The shaft 72 withattached thrust plates 74 and 76 rotates within a sleeve 78. The sleeve78 includes a portion having increased inside diameter to create an airspace 80. The bearing 70 includes external surface tension seals 82 andinternal surface tension seals 84. The external surface tension seals 82and internal surface tension seals 84 are connected by pressureequalization ports 86. The surface tension seals 82 and 84 and thepressure equalization ports 86 are filled with a lubricant. As with theembodiment of Fi. 1, the surface tension seals are created by diverging,axially extending surfaces.

The embodiment of FIG. 7 results in reduced evaporation rate of thelubricant from the seals. When the orientation of a bearing changes, theposition of the surface tension seals along the spin axis also changes.In the case in which the oil air interface moves into the bearing, afilm of oil is left on the region of the metal which was previouslycovered by the lubricant of the seal. This film of oil is then exposedto air and has a large amount of surface area compared to the seal oilair interface area. As a result of this increased surface area, theevaporation of the oil is increased and the life of the lubricant supplyis reduced.

When a bearing is not operating, the position of the seals is determinedby the pressure difference between the two sealed regions of the bearingwhich are connected together by the pressure equalization port orbalance tube. The internal fluid pressure difference is controlled bythe elevation difference between the two regions of the bearing and thespecific weight of the lubricant fluid. The external pressuredifferences due to variations in air pressure around the bearing areusually negligible. Thus seal position and the change in seal positionare controlled primarily by the elevation changes in the bearing.Splitting the lubricated regions of the bearing of FIG. 7 into twoseparate and shorter zones reduces the range of possible elevationdifferences and also the resulting range of seal position changes. Thisdesign thus reduces the wetted area of the bearing and the evaporationrate.

The bearing of FIG. 7 also provides higher moment stiffness. The higherstiffness results from the fact that the length of the bearing can bemade longer relative to the bearing of FIG. 1. Moment stiffness isproportional to the length of the bearing squared when all of the otherbearing characteristics are held constant. The bearing of FIG. 7 can belonger than the bearing of FIG. 1 because the seal areas are split intoseparate zones so that the central region of the bearing can belengthened without affecting the behavior of the seals.

Another advantage of the embodiment of FIG. 7 is faster thermaltransient response of the lubricant. It is desirable to have thelubricant come up to temperature as fast as possible during start up.When the lubricant oil is warm, it has a lower viscosity than when it iscold and thus the torque requirements are less when the oil is warm.Accordingly, when the oil can be made to heat up quickly, a shorterperiod of high load on the driving motor results which is very desirablefor some applications. The faster thermal response of the bearing ofFIG. 7 results from the reduction of oil volume in this bearing designand the resulting increase in bearing power to oil volume ratio.

FIG. 8 is yet another embodiment of the present invention. A bearing 100includes a spindle shaft 102 which has tapered portions 104 and 106.These tapered portions mate with tapered bearing shells 108 and 110which reside within a spindle housing or sleeve 112. The spaces betweenthe tapered shaft and tapered bearing shells are filled with a liquidlubricant. The lubricant is sealed by external surface tension orcapillary seals 114 and 116, and internal capillary seals 118 and 120.An equalization port 122 connects the seals 114 and 118, and anequalization port 124 connects the seals 116 and 120.

Because of the tapered surfaces, both radial and axial loads aresupported by the bearing. The spindle housing and shaft surfaces are asingle contiguous unit without any parting line. No O-ring seals arerequired since no secondary leakage is possible with the taperedarrangement. The tapered portions of the bearing shaft or the taperedbearing shell surfaces include herringbone patterns which generate a netliquid flow due to machining tolerances. This net liquid flow in thebearing is compensated for by a flow in the opposite direction throughthe equalization ports 122 and 124.

The bearing shells 108 and 110 have grooves on their outer surfaces.These bearing shells are shrink fitted into the spindle housing 112 andthe grooves cooperate with the housing 112 to create the equalizationports.

What is claimed is:
 1. Hydrodynamic bearing comprising:a rotatableshaft/thrust plate combination disposed within a sleeve forming a firstclearance space between the shaft and the sleeve and a second clearancespace between the thrust plate and the sleeve, the clearance spacesbeing filled with a liquid lubricant, the external faces of the thrustplates being exposed to air; the thrust plate and sleeve having axiallyextending surfaces which diverge toward the ends of the bearing in astraight taper of approximately 2° to form surface tension dynamicseals; the sleeve including pressure equalization ports connecting thefirst clearance space and the second clearance space.
 2. The bearing ofclaim 1 wherein the pressure equalization ports include axiallyextending passageways in communication with radially extendingpassageways.
 3. The bearing of claim 2 wherein the radially extendingpassageways are located near the center of the bearing.
 4. The bearingof claim 2 wherein the axially extending passageways are equally spacedaround the sleeve.
 5. The bearing of claim 1 including four equalizationports.
 6. The bearing of claim 1 further including static seals betweenthe shaft and the thrust plates.
 7. The bearing of claim 6 wherein thestatic seals are O-ring seals.
 8. The bearing of claim 1 furtherincluding relief patterns in opposed sleeve/thrust plate faces togenerate inwardly directed radial forces.
 9. Hydrodynamic bearingcomprising a cylindrical sleeve including a portion having a smallerinside diameter;a shaft including a portion having a diameter adapted toform a first clearance space with respect to the smaller diameterportion of the sleeve; a pair of thrust plates disposed on the shaftforming second clearance spaces with respect to radially extending facesof the smaller diameter portion of the sleeve, the external faces of thethrust plate being exposed to air, the clearance spaces being filledwith a lubricant; the smaller diameter portion of the sleeve includingplural axially extending passageways communicating with radiallyextending passageways interconnecting the first and second clearancespaces; and surface tension seals between the thrust plates and sleeve;the thrust plate and sleeve including axially extending surfaces whichdiverge toward the ends of the bearing in a straight taper ofapproximately 2° to form the surface tension seals.
 10. The bearing ofclaim 9 wherein opposed sleeve/thrust plate faces include reliefpatterns to generate inwardly directed radial forces.